Power delivery system having a continuously variable ratio transmission

ABSTRACT

A control system and method for a power delivery system, such as in an automotive vehicle, having an engine coupled to a continuously variable ratio transmission (CVT). Totally independent control of engine and transmission enable the engine to precisely follow a desired operating characteristic, such as the ideal operating line for low fuel consumption. CVT ratio is controlled as a function of commanded desired system performance (e.g., power or torque) and measured actual system performance, such as CVT torque output, while engine fuel requirements (e.g., throttle position) are strictly a function of measured engine speed. Fuel requirements are therefore precisely adjusted in accordance with the ideal characteristics of any load placed on the engine. Appropriate controls prevent anomalous engine and vehicle behavior, and allow for transient start-up from rest. CVT ratio is corrected to maintain target value operating characteristics and to ensure smooth and stable operations by a control circuit which controls the diameter of the CVT driver pulley while the diameter of the driven pulley is maintained constant.

BACKGROUND OF THE INVENTION

This invention relates to, and is an improvement upon, commonly assignedU.S. Pat. No. 4,459,878, issued July 17, 1984.

This invention relates to a power delivery system having a continuouslyvariable ratio transmission and, more particularly, to a control systemand a control method for such a system, such as might be used in anautomotive vehicle. In particular, the present invention relates to animproved control system and a control method for improving performanceand stability upon vehicle start-up and maintaining minimum fuelconsumption.

The quest for greater fuel economy of automotive vehicles has led tosignificant improvements in engine and transmission design and control.Continuously variable ratio transmissions (CVT) have shown particularpromise in this regard. It will be appreciated that at any given vehiclespeed, and for any needed propulsive force, a certain transmission ratiowill provide maximum fuel economy for a given engine. In addition, forany given vehicle speed, one transmission ratio will permit maximumacceleration with that engine. Since a CVT with the proper ratio rangecan provide any desired transmission ratio, it is obviously attractivefor automobiles from the standpoint of economy, low emissions andperformance. If the mechanical efficiency of the CVT is high and itsratio range is wide enough, it can even be possible to have both maximumeconomy and maximum performance in the same vehicle. Among the obviousbenefits are fully automatic operation, smooth, stepless and rapidresponse to driver demand, and quieter cruising.

Many different CVT configurations have been developed in the prior art.These include, for example, hydrostatic transmissions; rolling contacttraction drives; overrunning clutch designs; electrics; multi-speed gearboxes with slipping clutch; and V-belt traction drives. Of these theV-belt traction drives appear attractive for small to medium sizepassenger car applications because of their compactness, lightness andsimplicity of design. Basically, this type of CVT comprises a V-beltwhich interconnects a driver sheave and driven sheave, the diameters ofthe sheaves being variable to change the ratio of the CVT. Recentadvances in belt design have resulted in improved belt durability andlongevity. If sheave movement can be properly controlled so as to avoidundue stresses on the belt, it is expected that a very long belt lifecan be achieved.

Many control schemes have been devised for engine-CVT systems inattempts to maximize fuel economy. These have been based on empiricalanalyses of individual engine performance, and the realization that, forany desired power output, there is an optimum combination of enginespeed and torque which will result in mimimum fuel consumption. This isillustrated in FIG. 1.

FIG. 1 is a typical performance map of a four cylinder spark ignitionpassenger car engine having a displacement of approximately 2.5 liters.The map is a plot of engine torque T_(E) and brake horsepower BHP as afunction of engine speed N_(E). The dot-dash line near the top of themap is a plot of engine torque at full throttle. The series of curves insolid black lines are fuel consumption contours, indicating constantbrake specific fuel consumption (BSFC) in 1 b.M/BHP-hr. Minimum fuelconsumption occurs at a point designated by 0.4 pounds perhorsepower-hour. The series of dashed lines indicates power output ofthe engine. The ideal operating line for low fuel consumption isindicated by the heavy solid line f(N_(E)), this curve being a functionof engine speed. This minimum fuel consumption line is shown moreclearly in FIG. 4 as a function of engine speed and throttle opening.The ideal operating for low fuel consumption is purely a function ofengine characteristics and is optimal regardless of vehicle road speed.Other ideal operating lines may appear on the performance map, forexample, the ideal operating line for low emissions.

In a vehicle with a conventional, manually shifted gearbox, forwardspeed ratios usually are available in only four or five steps. Theoperating point of the engine on the performance map is determined bydrive shaft speed, power or torque commanded, and transmission gearratio. Since there are only a few gear ratios available in a typicaltransmission, the engine must be throttled much of the time. The enginemust therefore operate most the time at high BSFC values. In contrast, aCVT is able to vary its speed ratio continuously to allow the engine torun at wider throttle and lower BSFC values.

Perhaps the most difficult task demanded of a control system for anengine-CVT system is to maintain engine operation along the idealoperating line. This is due to the almost continuous transient nature ofoperation of an automotive vehicle, there being hardly ever a time whenroad load and commanded torque or power remain constant. Transientconditions usually are dealt with by a change in CVT ratio, engine speedand throttle. Prior art control systems, by their very nature, permit anexcursion of engine operation away from the ideal operating line beforereturning back to it at steady state. An example of such an excursion isshown in FIG. 1 by dashed line X-Y-Z. The result is that engineoperation approaches, but hardly ever is maintained on the idealoperating line. Two such prior art systems are illustrated in FIGS. 2and 3.

FIG. 2 schematically illustrates a system devised by Peter Stubbs forBritish Leyland. This system is described in greater detail in Stubbs,The Development of a Perbury Traction Transmission for Motor CarApplications, ASME Publication No. 80-C2/DET-59 (August, 1980). In thissystem, engine speed, throttle position and CVT ratio signals are allfed to a computer controller which has, in its memory, the engineoperating characteristic for minimum fuel consumption. The computercontroller generates, as a function of these variables, an enginecontrol signal for adjusting the position of the throttle, and a ratiorate signal which changes the ratio of the CVT. The throttle is underthe direct control of the vehicle accelerator pedal so that, while theengine control signal may vary the throttle position somewhat from thatcommanded by the driver, the throttle position still is primarily afunction of commanded power or torque.

FIG. 3 is a schematic representation of the system devised by Miyao forAisin Seiki. This system is described in greater detail in U.S. Pat. No.4,091,690. Here, as in the Stubbs system, engine throttle is primarily afunction of commanded power or torque by direct connection with theaccelerator pedal. The computer generates a ratio rate signal to changetorque and speed. Inherently sensed output torque also affects the CVTratio.

In these, as well as in virtually all other engine-CVT control systems,throttle position is controlled directly by the vehicle acceleratorpedal, or is a direct function of pedal position, as well as otherparameters. Engine and transmission control usually are directly relatedto one another. Such control schemes permit engine operation duringtransients to vary from the ideal operating line. Excursions away fromthe ideal operating line result in less than optimum engine operation(e.g., excessive fuel consumption, or excessive emissions), untileffective control is resumed by the system during steady stateoperation. As pointed out earlier, however, most vehicular operation istransient in nature, rather than steady state, so that substantially allengine operation occurs off the ideal operating line. This problem isparticularly accute when a vehicle first begins to move, from rest.Emissions calibrations must therefore be made in a substantial portionof the engine performance map. Most prior art control systems also mustbe specifically tailored to particular engines. This requires numerousspecially designed control systems for a fleet of differently poweredvehicles. In addition, most prior art control systems cannot compensatefor varying engine conditions, the result being vehicle driveabilitywhich varies with engine temperature, state of tune, age and altitude.Close duplication of conventional vehicle characteristics also is aproblem with prior art CVT control schemes.

SUMMARY OF THE INVENTION

It is therefore an object of the present invention to overcome theabove-noted disadvantages and deficiencies of the prior art by providingan engine-CVT control scheme which improves upon the scheme disclosed inU.S. Pat. No. 4,459,878 for substantially constantly maintaining engineoperation along the ideal operating line.

Another object of the invention is to provide such a control schemewhich yields substantially constant vehicle driveability as sensed bythe driver, irrespective of engine temperature, age, state of tune,altitude and other variables.

Another object of the invention is to provide such a control scheme in avehicle whose characteristic will remain the same irrespective of thetype of engine which is coupled to the CVT.

Another object of the invention is to provide such a control scheme in aCVT vehicle, which will enable the vehicle to perform almost in allrespects as a vehicle with a conventional transmission.

Another object of the invention is to provide a power delivery systemwhich maintains stability and drivability by smoothly changing a fuelfunction or throttle opening in response to engine operating speed orcommanded system performance and thus maintain minimum fuel consumptionupon clutch engagement.

A further object of the invention is to provide a correction to the CVTratio upon clutch engagement or at other times to thus maintainstability and driveability independent of throttle opening.

Another object of the invention is to greatly simplify calibration ofthe engine for emissions purposes.

As disclosed in U.S. Pat. No. 4,459,878, engine operation can readily bemaintained along the ideal operating line by providing for totallyindependent engine and transmission control. That is, the position ofthe engine throttle is totally independent of accelerator pedalposition. Throttle position and, hence, engine output torque simply is afunction of engine speed only, and that function may be any desiredrelationship, for example, the ideal operating line for low fuelconsumption, the ideal operating line for low emissions, or a compromiseideal operating line for low fuel consumption and low emissions. Torque,power or other desired performance parameters commanded by theaccelerator pedal controls CVT ratio, and engine speed is determined bythe load placed thereon, which is a function of road load and CVT ratio.Hence, throttle position is precisely adjusted in accordance with theideal function for any load placed on the engine. With appropriatelydesigned controls, which also are a part of this invention, anomalousengine and vehicle behavior, such as engine overspeed and underspeeedconditions, can be prevented, transient start-up from rest can beaccommodated, and the vehicle can be made to perform almost in allrespects just as a vehicle with a conventional automatic transmission.

For convenience, the invention is described throughout thisspecification in the context of an engine-CVT propulsion system for anautomotive vehicle. It is to be understood, however, that the principlesof the invention are equally applicable to any type of power deliverysystem, including but not limited to other vehicular systems usinginternal or external combustion engines of any design, or to stationarypower plants for driving compressors, generators or any other type ofmachinery. Where the term "throttle" is used, the term is understood toencompass any mechanism for controlling the delivery of fuel to theengine or other prime mover, be it a conventional carburetedspark-ignition engine wherein fuel flow varies with throttle butterflyposition, a fuel injected spark-ignition or diesel engine, a gasturbine, and so on. Where the term "clutch" is used, the term isunderstood to encompass any type of frictional torque converter or othercoupling used to connect and disconnect a driving and driven part of amechanism.

U.S. Pat. No. 4,459,878 discloses a method of controlling the operationof a power delivery system including a prime mover and a continuouslyvariable ratio transmission coupled to the prime mover for deliveringpower from the prime mover to an output shaft. The prime mover has fueldelivery means for delivering a variable quantity of fuel thereto, andthe power delivery system is controlled by command means for commandinga desired system performance parameter, such as output power or torquedelivered to the output shaft. The method includes the steps ofmeasuring the actual performance of the system, and controlling theratio of the transmission as a function of the commanded performanceparameter, the measured actual performance of the system, and a targetmimimum fuel consumption performance parameter. The speed of the primemover varies as a function of transmission ratio. A fuel function ispredetermined which defines desired fuel requirements for the primemover in relation to prime mover operating speed. The speed of the primemover is measured, and the fuel delivery means is controlled inaccordance with the fuel function so that the fuel delivered to theprime mover is determined only by the speed thereof.

In vehicular applications, during stationary and relatively slow vehicleoperation, the command means is temporarily operatively coupled to thefuel delivery means to provide positive driver control of the engine.The engine and transmission may be coupled by a clutch or similarcoupling device, such as a fluid coupling, which is disengaged when thevehicle is stationary and is partially engaged during slow vehicleoperation.

U.S. Pat. No. 4,459,878 also discloses a system for carrying out theabove-described method, and a power delivery system including the primemover, transmission and control system therefor. A method of controllingthe operation of the engine of an engine-driven vehicle duringstationary and relatively slow vehicle operation also is disclosed. Thevehicle engine is coupled to a CVT and has command means for commandinga desired output power or torque delivered to the output shaft, thedrive ratio of the transmission varying as a function of commanded poweror torque to thereby cause the speed of the engine to vary. The methodcomprises the steps of predetermining a fuel function defining desiredfuel requirements for the engine in relation to engine operating speed,measuring the speed of the engine, controlling the fuel delivery meansonly in accordance with the fuel function so that the fuel delivered tothe engine is determined only by the speed thereof, and operativelycoupling the command means to the fuel delivery means during stationaryand relatively slow vehicle operation. This coupling is dependent uponcomparison of a ratio signal, equal to the quotient of measured enginespeed and measured output shaft speed, to a predetermined slow operationratio, the command means being operatively coupled to the fuel deliverymeans when the ratio signal exceeds the slow operation ratio. A systemfor controlling the engine in accordance with this method also isincluded.

The present invention improves upon the invention disclosed in U.S. Pat.No. 4,459,878 and provides a method whereby engine rotational speed ismaintained at a minimum fuel consumption level by correcting the CVTratio (R) independently of throttle opening degree, thereby to obtainstable running conditions. After clutch engagement, fuel consumption canbe reduced as well by determining a desired throttle opening degree inresponse to the engine rotational speed as usual.

The present invention provides a method of controlling the operation ofa power delivery system including a prime mover and a continuouslyvariable ratio transmission coupled to the prime mover for deliveringpower from the prime mover to an output shaft. The prime mover has afuel delivery means for delivering a variable quantity of fuel to theprime mover. The power delivery system of the present invention iscontrolled by a command element, such as the accelerator pedal of avehicle, which commands a desired system performance.

The transmission of the present invention has a variable driver side forreceiving power from the prime mover and a variable driven side fordelivering power to the output shaft. The speed of the prime movervaries as a function of the transmission ratio.

According to the present invention, the method of controlling theoperation of a power delivery system measures the actual performance ofthe system and the speed of the prime mover. The measured speed of theprime mover is compared to a target prime mover speed value. The ratioof the transmission is controlled by controlling one side, e.g., thedriven side, of the transmission as a function of the desired systemperformance commanded by the accelerator pedal or other command means,measured output shaft speed, and measured engine speed. The other sideof the transmission, e.g., the driver side, is controlled as a firstcontrol function of measured actual system performance when the measuredprime mover speed equals the target prime mover speed value or as asecond control function of commanded system performance when themeasured prime mover speed exceeds the target speed value, whereby theload on the prime mover is adjusted to converge the prime mover speedwith the target prime mover speed.

The first control function preferably controls the transmission as afunction of measured system output torque (T_(O)), measured engine speed(N_(e)) and the ratio signal equal to the quotient of measured primemover speed and measured output shaft speed (R).

Preferably, the present invention includes a clutch which is disengagedwhen the vehicle is stationary, and is partially engaged during slowvehicle operation. In order to prove improved performance andoperability, the present invention further predetermines a first fuelfunction corresponding to desired fuel requirements for the prime moverin relation to its operating speed. The second fuel function ispredetermined corresponding to desired fuel requirements for the primemover in relation to commanded system performance. The fuel deliverymeans are controlled in accordance with the first fuel function when theclutch is fully engaged and the measured prime mover speed equals thetarget prime mover speed. The fuel delivery system is controlled inaccordance with the second fuel function when the clutch is less thanfully engaged or when the measured prime mover speed does not equal thetarget prime mover speed value.

In order to measure the condition of clutch slippage in the presentinvention, the speed of the output shaft is measured. A ratio signal isthen generated equal to the quotient of measured prime mover speed andmeasured output shaft speed. This ratio signal is compared to apredetermined slow operation ratio. Clutch slippage occurs when theratio signal is greater than the slow operation ratio.

BRIEF DESCRIPTION OF THE DRAWINGS

FIGS. 1-3 correspond to identical figures in U.S. Pat. No. 4,459,878;FIGS. 5-10 correspond to FIGS. 4-9 of said patent, respectively; FIGS.12 and 13 correspond to FIGS. 10 and 11 of said patent, respectively;and FIGS. 4 and 11 are directed specifically to the present invention.

The novel features of the invention are set with particularity in theappended claims, but the invention will be understood more fully andclearly from the following detailed description of the invention as setforth in the accompanying drawings, in which:

FIG. 1 is the performance map of a typical four cylinder passenger carengine having a displacement of approximately 2.5 liters;

FIGS. 2 and 3 illustrate two forms of prior art engine-CVT controlschemes;

FIG. 4 is a graph of an improved control scheme in accordance with thepresent invention and shows a minimum fuel consumption line plottedagainst the degree of throttle opening as a function of enginerotational speed.

FIG. 5 is a schematic illustration showing the functional relationshipsof the components of an engine-CVT control scheme used with theinvention;

FIG. 6 is a schematic illustration showing the entire control systemused with the invention and its relationship to the CVT sheave and beltdrive, and the vehicle starting clutch;

FIG. 7 is a graph which shows the forces applied to the driver anddriven sheaves of the CVT as a function of transmission ratio.

FIGS. 8 through 10 together schematically represent the entireengine-CVT control scheme used with the invention, the figures beinginterrelated as indicated therein by lines A-B and C-D;

FIG. 8 primarily relates to the engine control circuit;

FIG. 9 primarily relates to the starting clutch control circuit;

FIG. 10 primarily relates to the sheave pressure generators;

FIG. 11 is a schematic illustration showing the control system accordingto the present invention for controlling clutch engagement as a functionof engine rotational speed in order to maintain minimum fuelconsumption.

FIG. 12 illustrates a modification of the pressure generator for thedriven sheave illustrated in FIG. 10; and

FIG. 13 is a graphical representation of the operation of an Engine-CVTsystem used with the control scheme of the invention.

DETAILED DESCRIPTION

FIG. 5 illustrates the general functional relationships of thecomponents of a power delivery system used with the present invention.An engine 10 is drivingly coupled to a continuously variable ratiotransmission (CVT) 14 through a clutch or fluid coupling (not shown).Fuel is fed to engine 10 by a fuel delivery means 12, which may be thethrottle and fuel jets of a conventional carburetor, a fuel injectionsystem or the like. CVT 14 may be one of the many types of continuouslyvariable ratio transmissions discussed above in connection with theprior art, although the V-belt traction drive type of CVT is preferred.Output shaft 16 delivers power and torque from the engine and CVT. Theratio of the CVT is set by a CVT ratio controller 17, which generates arate of change of ratio signal kR as a function of output torque T_(O)measured by torque sensor 19 and commanded power or torque α commandedby accelerator pedal 18. Other parameters indicative of engine-CVTsystem performance may be used by ratio controller 17 to effect a changeof CVT ratio in a similar manner. For example, rather than using desiredoutput power or torque and measured actual output torque, commanded andmeasured vehicle acceleration, output shaft acceleration, or otherparameters could be used. In this preferred embodiment, however, CVTratio is strictly a function of commanded power or torque and measuredoutput torque, and is completely independent of engine operation. Enginecontrol, on the other hand, is provided by an engine controller 100 withadjusts fuel delivery means 12 in accordance with measured engine speedN_(E). This relationship may desirably be the ideal engine operatingline for low fuel consumption, the ideal operating line for lowemissions, a compromise of the two, or any other desired engineoperating characteristic.

FIG. 6 schematically illustrates the entire control system in greaterdetail. The particular type of CVT illustrated in FIG. 5 is the variablediameter pulley, V-belt traction drive type having a driven sheave 20connected to output shaft 16 and a driver sheave 30 which is coupled toengine 10. Belt 15 engages the grooves in the sheaves 20 and 30 totransmit motive power therebetween. Sheaves 20 and 30 are hydraulicallyactuated by pressurized fluid to vary the driving diameters. Sheave 20has an axially fixed portion 22 and an axially movable portion 24.Pressurized fluid in a fluid chamber 26 behind movable portion 24provides the axial force required to maintain portions 22 and 24 at afixed distance from one another (i.e., to hold the driving diameter ofscheave 20 constant), and to move portion 24 toward or away from portion22 to vary the driving diameter. Similarly, sheave 30 has an axiallyfixed portion 32 and a movable portion 34 which is under the influenceof fluid pressure in chamber 36. Proper pressures in chambers 26 and 36to keep belt 15 under proper tension are maintained by the controlsystem, as described below.

The position of throttle (fuel delivery means) 12 is controlled by athrottle servo 13 which receives signals from engine control circuit100. During certain transient operations (described below) fuel deliverymay be diminished by a fuel diminishing value 11, or fuel delivery maybe suspended completely by a fuel suspension mechanism 9. The fueldiminishing and suspension functions may be performed, for example, by asingle solenoid valve operable in variable modes. Engine control circuit100 is responsive to inputs from the accelerator pedal, engine speed(N_(E)), a manual override switch which permits operation in theautomatic or manual mode, and a start/neutral switch (S/N) which insuresthat the vehicle will remain stationary when the engine is started.

Fluid pressure for activating the driven sheave is provided by a sheavepressure generator 200 which acts through a pressure servo controller250 and a fluid distribution circuit 500. Similarly, fluid pressure foractivating the driver sheave 30 is provided by sheave pressure generator300 acting through a servo controller 350 and fluid distribution circuit500. Pressure generator 200 is responsive to inputs of engine speedN_(E), accelerator position α, drive shaft speed N_(DS) measured by asensor associated with drive shaft 16, and CVT ration R. Ratio R isgenerated by CVT ratio circuit 600 and is the quotient of engine speedN_(E) divided by drive shaft speed N_(DS).

A starting clutch 40 is provided which couples engine 10 and CVT 14.Clutch 40 is disengaged when the vehicle is stationary, and is partiallyengaged during slow vehicle operation, gradually approaching fullengagement, which occurs as described below at a predetermined point ofoperation. Starting clutch 40 is controlled by a control circuit 400which is responsive to accelerator pedal position α, engine speed N_(E),and the auto/manual switch, through servo controller 450 and fluiddistribution circuit 500. Particularly during the transient condition ofstart-up, it is both desirable and difficult to maintain performanceparameters along the minimum fuel consumption line, as shown in FIG. 4.In accordance with the improvement of the present invention and as shownin FIG. 4, if the engine speed N_(E) is greater than a target enginespeed (N_(et)) along the minimum fuel consumption line for a constantthrottle opening position (point A), a comparing circuit, shown in FIG.11 and described in detail below, changes the CVT ratio to one having alarger load so that the actual engine speed N_(e) is reduced to thetarget engine speed N_(et) on the minimum fuel consumption line, asshown in point C.

FIGS. 8, 9 and 10 schematically illustrate in greater detail thefunctional relationships of several of the components illustrated inFIG. 6. FIG. 8 is primarily directed to the engine control circuit 100.A key element of control circuit 100 is function generator 102, whichmay generate a function representative of any desired engine operatingcharacteristic. For this embodiment the function φ is chosen as theideal operating line for low fuel consumption. φ represents throttleangle which is proportional to desired engine output torque. FIG. 1graphically illustrates this function as f(N_(E)). The value of thefunction produced by generator 102 is fed directly to throttle servo 13via amplifier 104. In the event the automatic control system isdisabled, it is possible to switch to a manual mode through mode switch106. In the manual mode, accelerator postion α is directly communicatedto throttle servo 13 via amplifier 104. The start/neutral switch S/Nalso operates through mode switch 106.

A fuel suspension comparator 108 provides backup engine overspeedcontrol, which may tend to occur upon vigorous acceleration, if there isa malfunction in the control system. Comparator 108 compares enginespeed N_(E) to the maximum permissible engine speed, for example 5500rpm. If N_(E) is greater than 6000 rpm, fuel suspension mechanism 9 isactivated to suspend delivery of fuel to engine 10. Fuel suspensionmechanism 9 may be, for example, a solenoid cutoff valve.

Another engine speed control is provided to counteract the inherenttendency of the vehicle to speed up when the accelerator pedal isreleased. This phenomenon occurs because the vehicle inertia becomescoupled to the interia of a relatively unthrottled engine through atransmission whose ratio is changing towards overdrive. This undesirabletendency is even more pronounced when the accelerator pedal is releasedsuddenly and completely. This anomalous behavior is prevented byreducing fuel flow to the engine when pressure on the accelerator pedalis relieved, the reduction of fuel flow being proportional to the rateat which pedal position decreases (-α), and by reducing fuel flow evenfurther when the accelerator pedal position α drops to below 3.6% offull excursion. To accomplish this control, a pulse width modulator 110controls fuel diminishing valve 11, the duty cycle (i.e., the percentageof the pulse cycle during which the fuel diminishing valve is held open)of modulator 110 being inversely proportional to the rate at which pedalposition α decreases (-α). -α is derived from a differentiator 112 onlyif α is less than zero. In addition, a fuel diminishing comparator 114reduces the duty cycle of modulator 110 to zero or nearly to zero whenpedal position drops to below 3.6%.

FIG. 9 relates primarily to the starting clutch control circuit 400. Itwill be appreciated that some type of coupling must be provided betweenthe engine and the CVT in order to permit the engine to idle while thevehicle is stationary. A fluid coupling could be used, but themechanical losses inherent in such a device are antithetical to thedesired objective of maximizing fuel economy. A torque converter with alock-up clutch would be an improvement, but a mechanical clutch ispreferred, and one which is hydraulically actuated would be well suitedfor this purpose. The object here, as in the conventional automobile, isto totally disengage the clutch when the vehicle is stationary, and togradually engage it to begin vehicle movement and progressively engagethe clutch furhter as the vehicle speed increases. To this end themeasured transmission ratio R (which is computed as the quotient ofengine speed N_(E) and drive shaft speed N_(DS) by ratio circuit 600) isfed to a comparator 402. Comparator 402 closes switch 404 when R exceeds4.7 to deliver the signal from amplifier 406 to throttle servo 13 viaamplifier 104. This signal is equal to α-N_(E) ', where N_(E) ' is afunction produced by generator 408 equal to K (N_(E) -1000 rpm). Thus,the accelerator pedal 118 is coupled directly to throttle 12 in avariable way defined by α-N_(E) '. The constant K is selected such thatengine speed cannot exceed 2500 rpm if the clutch is not fully engaged.This direct coupling of accelerator pedal to throttle allows an input tobe provided to the system to initiate movement of the vehicle from astationary position.

Comparator 402 also closes switch 410 to transmit pedal positiondirectly to the clutch pressure servo controller 450. Hence, the degreeof engagement of clutch 40 is proportional to pedal position up to thepoint where ratio R equals 4.7. During this period the degree of directcontrol of the accelerator pedal over throttle 12 diminishes as enginespeed increases in accordance with the above-described relationship.

When ratio R drops below 4.7, switches 404 and 410 open, and comparator411 closes switch 412 to deliver maximum pressure to the clutch servocontroller 450. Maximum pressure causes full engagement of the clutch.As the vehicle accelerates beyond this point, it is under totalautomatic control.

It can be seen that if no start/neutral S/N switch were provided, anydepression of accelerator pedal 18 upon startup would cause engagementof clutch 40 and a forward lurch of the vehicle. The S/N switchtherefore effectively disables the effect of α on clutch 40 t0 permitsafe startup.

FIG. 10 relates primarily to the sheave pressure generator for thedriven sheave 200 and the sheave pressure generator for the driversheave 300. Pressure generator 200 includes circuitry which changes thetransmission ration to increase the load on the engine if the enginetends to exceed the maximum operating speed of 5500 rpm (N_(MAX)). Alsoprovided is circuitry for changing the transmission ratio to decreasethe load on the engine should the engine speed tend to decrease belowthe idle speed of 1000 rpm (N_(MIN)). This is accomplished by means ofsumming amplifiers 230, 232 and clipping circuits 234, 236. Summingamplifier 232 and clipping circuit 236 act to reduce pressure on thedriven sheave 200 to increase the load on the engine. Amplifier 232receives N_(E), applied to its negative input terminal, and N_(MAX),applied to its positive input terminal, and produces a summed outputsignal N_(MAX) -N_(E). This summed output is applied to clipping circuit236 which is a non-linear device having the characteristic shown in FIG.9. This device can be, for example, a reverse biased diode whichproduces a negative substantially linear output for negative excursionsof its input signal and a zero input for positive excursions.

Consequently, if N_(E) exceeds N_(MAX), the input signal applied tocircuit 236 will be negative, thereby resulting in a negative outputsignal. This negative output signal is then applied to summing amplifier210 to reduce the value of its summed output signal in proportion to theamount N_(E) exceeds N_(MAX). As a result, the pressure on driven sheave200 will be proportionally decreased. On the other hand, if N_(E) isless than N_(MAX), the input signal applied to clipping circuit 236 willbe positive resulting in a zero output signal applied to amplifier 210.Such an output signal has no affect on the summed output signal ofamplifier 210; thus, no change in the signal supplied to the drivenservo-controller 250 is produced.

Summing amplifier 230 and clipping circuit 234 act to increase pressureon the driven sheave 200 to decrease the load on the engine. Amplifier230 receives N_(E), applied to its negative input terminal, and N_(MIN),applied to its positive input terminal, and produces a summed outputsignal N_(MIN) -N_(E). This summed output is applied to a clippingcircuit 234 similar to circuit 236. Circuit 234, however, has anonlinear transfer characteristic which produces a positivesubstantially linear output for positive excursions of its input signaland a zero output for negative excursions. Circuit 234 can be, forexample, a forward biased diode. If N_(E) falls below N_(MIN), the inputsignal applied to clipping circuit 234 will be positive, therebyresulting in a positive output signal. This positive output signal isthen applied to summing amplifier 210 to increase the value of itssummed output signal in proportion to the amount N_(E) is less thanN_(MIN). As a result, the pressure on driven sheave 200 will beproportionally increased. On the other hand, if N_(E) is greater thanN_(MIN), then a zero output signal will be produced by circuit 234 whichhas no affect on the summed signal applied to servo-controller 250.

Pressure generator 200 also includes circuitry for adjusting thesensitivity of accelerator pedal 18, depending on vehicle speed, to moreclosely simulate the "feel" of a conventional vehicle. This is requiredbecause of the inherent operating characteristics of the engine and CVT.That is, at higher vehicle speeds, the torque produced by the engineremains fairly high and constant (see FIG. 1). In the conventionalvehicle the remaining small percentage of torque which can be extractedfrom the engine is delivered to the rear wheels through a transmissionin high gear with a fixed, very low reduction ratio. Vehicleacceleration is therefore fairly insensitive to accelerator pedalmovement at high speeds. In a CVT equipped vehicle, however, depressionof the accelerator pedal even at high vehicle speeds results in anincreased reduction ratio and an associasted multiplication of torque inexcess of that provided in the conventional vehicle. Thus, if onlydirect accelerator pedal position a were used to control CVT ratio athigher vehicle speeds, vehicle response would be extremely sensitive toaccelerator pedal movement. The sensitivity of the accelerator pedal 18must therefore be dulled at higher vehicle speeds.

Pedal sensitivity is controlled by two comparators 212, 214. As long asvehicle speed is below a threshold equivalent of drive shaft speedN_(DS) equal to or less than 1173 rpm, switch 216 remains closed todeliver the a signal directly to amplifier 210. This effectively istorque control. When drive shaft speed N_(DS) exceeds 1173 rpm, switch216 opens and switch 218 is closed so that a pedal position signalequivalent to α divided by N_(DS) (provided by divider 220) is deliveredto amplifier 210. This effectively is power control. In this way, theeffect of any movement of accelerator pedal 18 in the higher speedranges is diminished so as to more closely simulated the pedal responseof a conventional automobile.

FIG. 11 shows an improved control circuit in accordance with the presentinvention for controlling the CVT ratio. In FIG. 11, control circuits 54and 55 control the operation of driver side pulley 30 and driven sidepulley 20, respectively. Control circuits 54 and 55 provide a standarddrive circuit with an operational output signal to actuator 53 of driverside pulley 30 and actuator 56 of driven side pulley 20. Control circuit54 provides a first control function of driver side pulley 30 as afunction of measured actual system performance when measured enginespeed (N_(e)) equals target engine speed (N_(et)). The preferred systemperformance parameters to be measured, as shown in block 54, includeoutput torque (To), measured engine speed (N_(e)), and the transmissionratio (R).

A selecting circuit 60 selectively switches over switches 50 and 62 whenit receives a signal that the CVT ratio R is less than or equal to 4.7,and that the actual engine speed N_(e) is equal to the target enginespeed N_(et) and thus agrees to the rotational speed of the minimum fuelconsumption corresponding to the throttle opening.

When the logic selecting circuit 60 receives the signal that clutch 40is fully engaged (i.e., R≦4.7) and when the actual engine rotationalspeed N_(e) equals the target rotational speed N_(et) on the minimumfuel consumption line (FIG. 4), switch 62 is switched to control circuit54 for driver side pulley 30.

When R is less than or equal to R₀ (4.7) and thus the clutch 40 is fullyengaged, and when the actual engine rotational speed N_(E) is equal tothe target engine rotational speed N_(et) control circuit 54 isconnected to actuator 53 by switch 62 which is placed to position B bycontrol circuit 60. As a result, the driven sheave 30 is controlled inaccordance with the measured CVT ratio (R), measured output torque, andmeasured engine speed. Throttle opening φ is controlled by actuator 51which receives the control signal via switch 50. When switch 50 is inposition B throttle control circuit 65 provides the fuel deliverycontrol in accordance with a first fuel function (φ), which is afunction measured engine speed N_(e), as previously discussed.

When R is greater than 4.7, that is when the clutch is slipping, andwhen the actual engine speed N_(E) is not equal to the target enginespeed N_(et), control circuit 60 places switches 62 and 64 and switch 50in position A and actuator 53 is thus provided its control signal by asecond control function, shown in block 52 and a second fuel function(φ') or throttle signal f(α) is provided to actuator 51. During thistime, fuel delivery and the driver sheave 30 are both controlled inaccordance with command power or torque.

Reference 63 indicates a select changeover means for correcting the CVTratio through actutor 53 by switching over the switch 64 from position Bto position A, as shown in FIG. 11, thereby to operate control circuit52 for driver side pulley 30 when clutch 40 is not completely engagedand when the engine rotational speed N_(e) is not equal to the targetengine rotational speed N_(et) for a particular throttle opening asshown in the graph on FIG. 4. The modification to the CVT ratio may becarried out by changing the diameter of the driver side pulley 30 asdescribed above.

Thus, using the control system described and shown in FIG. 11, if, uponclutch engagement, the engine rotational speed is off the minimum fuelconsumption line, the CVT ratio can be corrected independently ofthrottle opening by controlling the diameter of the CVT driver pulleywhile the diameter of the driven pulley is maintained constant. Thiswill produce more stable operating characteristics.

FIG. 12 shows a modification of the sheave pressure generator 200,wherein accelerator sensitivity is controlled as a function of ratio R.Comparator 212' closes switch 216' to connect the accelerator pedalposition signal a directly to amplifier 210 when ratio R equals orexceeds 3. The comparator 214' closes switch 218' to feed a dulledsignal to amplifier 210 from divider 220' when ratio R is below 3.

The control of transmission ratio described above actually is a ratiorate control, R. That is, the greater the magnitude of the increase (ordecrease) in fluid pressure on driven sheave 20 commanded by acceleratorpedal 18, the more rapid the change of sheave diameters will be. Thus,for example, a rapid depression of accelerator pedal 18 will result in arapid change of CVT ratio and quick acceleration. This, of course,closely simulates the characteristics of a conventional vehicle.

The instant invention involves, in part, the recognition that control ofthe ratio rate R of the CVT, rather than merely the CVT ratio, yieldsimproved CVT control. This improved control is explained by reference tothe following derived vehicle performance equation: ##EQU1## whereI_(EQ) =I_(CDS) +R² I_(E),

R is the ratio rate of the transmission,

R is the ratio of the transmission,

I_(E) is engine inertia,

N_(E) is engine speed,

T_(E) is engine torque,

T_(RL) is road load torque reflected to the drive shaft, and includestires, final drive and axle losses,

T_(loss) is transmission loss,

I_(CDS) is car inertia reflected to the drive shaft, and

N_(DS) is vehicle acceleration measured at the drive shaft.

It is clear that the acceleration of the vehicle N_(DS) is dependentprimarily upon control of any one or more of these variables such as,for example, T_(E), R or R. Generally, conventional vehicle systems varythe transmission ratio R and engine output torque T_(E) to provide therequired transmission and vehicle control. By controlling R, however, itis difficult to constantly maintain engine torque and speed along theideal operating line. This is due to the fact that each time R isvaried, the load on the engine is changed which, in turn, affects theengine's output torque and vehicle acceleration.

Attempts to simultaneously change the engine torque and speed to forceengine operation back on the ideal line have necessitated very complexcontrol systems, since control is dependent on several variables of theperformance system. For example, these systems must necessarily performthe complicated task of calculating the necessary target throttleposition and CVT ratio R to force engine operation back on the idealline. These systems also require the calculation of ratio rate R so thatthe rate in changing the ratio to the target value does not result inundesirable vehicle dynamics. For example, if R is selected to beexcessive then an undesirable deceleration of the vehicle will occurbefore the vehicle can accelerate. This phenomenon results from thenegative sign of the R term in the above performance equation.

This invention, however, recognizes that R can easily be sensed andcontrolled without causing the other variables to adversely affectengine performance. This is accomplished by separating the enginecontrol from the transmission control so that engine torque and speedare fixed along the ideal engine operating line. As a result ofcontrolling R no adverse effect on the other dependent variables occurs.In particular, changing R alone, with its concomitant change on R, willnot force engine operation off the ideal operating line since enginespeed and torque are determined solely by the fuel function f(N_(E)). Asa result, vehicle acceleration N_(DS) and output torque T_(O) arecontrolled solely by ratio rate R, rather than by the other variables ofthe performance system.

It has been discovered in accordance with this invention that rate ofchange of ratio (R) is closely approximated by the followingrelationships:

    k R=α-T.sub.O

(for low speeds: torque control) and

    k R =α/(k'N.sub.DS)-T.sub.O

(for high speeds: power control)

In the V-belt traction drive CVT of the preferred embodiment of theinvention, the comparison of accelerator pedal position α and outputtorque T_(O) occurs inherently in the belt and pulley components toeffect a ratio change at a rate R. Other types of CVTs may requiredifferent control elements to effect this relationship. As statedearlier, however, other parameters indicative of system performance maybe used to effect a ratio change at a rate R, where R is proportional tothe difference between the desired performance parameter and the actualmeasured performance parameter.

The above described control scheme of the invention is graphicallyillustrated in FIG. 13. FIG. 13 is a plot of engine speed N_(E) as afunction of vehicle speed or drive shaft speed N_(DS). The minimum andmaximum CVT ratios are illustrated by the straight line emanating fromthe origin of the graph. The idle speed (N_(MIN) =1000 rpm) is indicatedby a lower horizontal line, while the maximum permissible engine speed(N_(MAX) =5500 rpm) is indicated by an upper horizontal line. Themaximum vehicle speed is defined by a vertical line at the right handedge of the graph.

The graph of FIG. 13 is divided into a number of discrete operatingregions. "A" designates the normal region of operation of the engine-CVTsystem. Region "A" is bounded by the line of maximum CVT ratio, the lineof maximum engine speed, the line of maximum vehicle speed, the line ofminimum CVT ratio and the idle speed line. During operation of thesystem in region "A", clutch 40 is fully engaged and throttle positionis wholly a function of engine speed in accordance with the fuelfunction f(N_(E)). Operation to the left of the dashed vertical lineindicating a drive shaft speed of 1173 rpm is under torque control,while operation to the right of this line is under power control (seethe above two equations, and the accelerator pedal sensitivity circuitryillustrated in FIGS. 10 and 12. Region "B" is the region of start-upcontrol, that is, the operation of the engine-CVT system during slowvehicle operation when clutch 40 is only partially engaged. The controlfor this operation (400) is illustrated in FIG. 9.

Operation of the engine-CVT system in the remaining three regions "C","D", and "E" is effectively prevented by the above described controlsystem. That is, operation in region "C" is prevented by the physicallimitation of minimum CVT ratio, and by the fuel diminishing circuitscomprising fuel diminishing valve 11, pulse width modulator 110,differentiator 112 and fuel diminishing comparator 114 of engine controlcircuit 100 (FIG. 8). Region "D" is the region of overspeed control,governed by the fuel suspension mechanism 9 and fuel suspensioncomparator 108 of engine control circuit 100 (FIG. 8) and by amplifier232 and clipping circuit 236 of sheave pressure generator 200 (FIG. 10).Region "E" is the region of engine idle control which is governed byamplifier 230 and clipping circuit 234 of sheave pressure generator 200(FIG. 10).

Also shown on the graph of FIG. 13 is a load line which indicates theengine speed required to maintain any constant vehicle speed along alevel road. The term "load" includes road load, final drive losses andthe like, and represents the actual load on the engine-CVT system. Inorder for the control scheme of the invention to function only inaccordance with the fuel function so as to maintain engine operationalong the ideal operating line, it is desirable that the CVT ratio rangeinclude substantially all ratios required to maintain constant vehiclespeed for any normally encountered load. That is, the minimum CVT ratiopreferably is smaller than that required to maintain constant vehiclespeed along a level road, and the maximum CVT ratio preferably isgreater than that required to maintain constant vehicle speed up thesteepest grade which one might expect to encounter. This relationship isgraphically illustrated by the physical location of the load line in thegraph of FIG. 13 above the line of minimum CVT ratio in region "A". Allother load lines should lie below the line of maximum CVT ratio. Adesirable CVT ratio range for accomplishing this is approximately 11:1with, for example, a maximum CVT ratio of 22:1 (total vehicle ratio,including final drive ratio), and a minimum CVT ratio of 2:1. Atransmission having such a wide ratio range is dislcosed in commonlyassigned application Ser. No. 290,293, filed Aug. 5, 1981. Of course, aCVT having smaller ratio range would be operable, but would not have asmuch flexibility as one with a wider range.

Referring to FIG. 7, the mechanics of a change in CVT ratio now will bedescribed with reference to the axial forces produced by the pressurizedfluid in chambers 26 and 36. The lower curve in FIG. 7 is a plot ofsteady state axial force on movable portion 24 of driven sheave 20 as afunction of CVT ratio. Similarly, the upper curve is a plot of steadystate axial force tending to resist inward movement of movable portion34 as a function of CVT ratio. As described below, when for example in asignal is generated to increase the ratio of the CVT from 1.0 toapproximately 1.7, the fluid pressure in chamber 26 is increased toraise the axial force from approximately 175 kg. to, ultimately,approximately 270 kg. Movable portion 24 does not move instantaneously,however, due to the inertia of the system. Accordingly, the curve whichrepresents the transient change taking place in sheave 20 is defined bymovement from point A to point B at a constant ratio of 1.0, and then topoint C where equilibrium is reached. Correspondingly, an increase inpressure in chamber 36 of driver sheave 30 results in an increase inaxial force on movable portion 34 of sheave 30 from approximately 315kg. (point D) to approximatley 380 kg. (equilibrium point E). Despitethis increase in axial force, the increased tension on belt 15occasioned by expansion of the diameter of sheave 20 forces the twoportions 32, 34 of sheave 30 apart so that sheave 30 has a smallerdriving diameter. Driver sheave 30, therefore, follows in a controlledmanner any changes occurring to driven sheave 20.

Sheave pressure generator 300 generates a pressure appropriate fordriver sheave 30 as a function of ratio R and measured output torqueT_(O). This function has been found to satisfactorily tension belt 15,without undue stress and effect a smooth change of ratio. An example ofa function suitable for this purpose is as follows:

    P.sub.DR =K.sub.1 =(K.sub.2 /R+K.sub.3) T.sub.O

where P_(DR) is the fluid pressure chamber 36 of driver sheave 30, andK₁, K₂ and K₃ are appropriately selected constants. The above-describedcontrol scheme quite simply and effectively accomplises its primaryobjective of maintaining engine operation along the ideal operatingline, for example, that of minimum fuel consumption. Transmissioncontrol requires output torque and accelerator pedal position sensing,while engine control requires only engine speed sensing. The specificparameter values set forth in the preferred embodiment described aboveare in no way intended to limit the scope of the invention, it beingapparent that these parameters will vary in accordance with engine,transmission and vehicle design, and desired behavior and performance.Numerous modifications of the invention will be apparent to thoseskilled in the art without departing from the true spirit and scope ofthe invention which is defined by the appended claims.

I claim:
 1. A method of controlling the operation of a power deliverysystem including a prime mover and a continuously variable ratiotransmission coupled to said prime mover for delivering power from saidprime mover to an output shaft, said prime mover having fuel deliverymeans for delivering a variable quantity of fuel thereto, said powerdelivery system being controlled by command means for commanding adesired system performance, and said transmission having a variabledriver-side receiving power from said prime mover and a variabledriven-side delivering power to said output shaft, the speed of saidprime mover varying as a function of transmission ratio, the methodcomprising the steps of:measuring the actual performance of the powerdelivery system; measuring the speed of said prime mover; comparingmeasured prime mover speed to a target prime mover speed value; andcontrolling the ratio of said transmission by controlling one said ofsaid transmission as a function of the desired system performancecommanded by said command means, and controlling the other side of saidtransmission as a first control function of said measured actual systemperformance when said measured prime mover speed equals said targetprime mover speed value, and as a second control function of commandedsystem performance when said measured prime mover speed exceeds saidtarget speed value, whereby the load on said prime mover is adjusted toconverge said prime mover speed with said target prime mover speed. 2.The method of claim 1 wherein said power delivery system includes aclutch which is disengaged when said vehicle is stationary, and ispartially engaged during slow vehicle operation, further comprising thesteps of:predetermining a first fuel function defining desired fuelrequirements for said prime mover in relation to prime mover operatingspeed; predetermining a second fuel function defining desired fuelrequirements for said prime mover in relation to command systemperformance; controlling said fuel delivery means in accordance withsaid first fuel function when said clutch is fully engaged and saidmeasured prime mover speed equals said target prime mover speed value;and controlling said fuel delivery means in accordance with said secondfuel function when said clutch is less than fully engaged or saidmeasured prime mover speed does not equal said target prime mover speedvalue.
 3. The method of claim 1 wherein the condition of clutch slippageis determined by the steps of:measuring the speed of the output shaft;generating a ratio signal equal to the quotient of measured prime moverspeed and measured output shaft speed; and comparing said ratio signalto a predetermined slow operation ratio, clutch slippage occurring whensaid ratio signal is greater than said slow operation ratio.
 4. Themethod of claim 3 wherein said first control function controls saidtransmission as a function of measured system output torque, measuredprime mover speed and said ratio signal, and said second controlfunction controls said transmission as a function of commanded systemperformance.
 5. The method of claim 4 wherein said one side of saidtransmission is controlled as a function of command system performance,measured output shaft speed and measured prime mover speed.
 6. Themethod of claim 2 wherein said prime mover is an internal combustionengine.
 7. The method of claim 2 wherein said fuel delivery meanscomprises a throttle.
 8. The method of claim 2 wherein said prime moveris the engine of an engine-driven vehicle.
 9. A power delivery systemfor a power driven device comprising:a prime mover; a continuouslyvariable ratio transmission coupled to said prime mover; coupling meansfor coupling said prime mover and said transmission; an output shaftcoupled to said transmission for receiving power from said prime moverthrough said transmission; fuel delivery system means for delivering avariable quantity of fuel to said prime mover; command means forcommanding a desired power delivery system performance; saidtransmission having a variable driver side receiving power from saidprime mover and a variable driven side delivering power to said outputshaft, the speed of said prime mover varying as a function oftransmission ratio; actual system performance measuring means formeasuring the actual performance of the power delivery system; speedmeasuring means for measuring the speed of said prime mover; means forcomparing the measured prime mover speed to a target value speed; ratiocontrol means operatively coupled to said command means and said actualsystem performance measuring means for controlling the ratio of saidtransmission by controlling one side of said transmission as a functionof the desired system performance commanded by said command means andcontrolling the other side of said transmission as a first controlfunction of the measured actual system performance when said measuredprime mover speed equals said target prime mover speed value, and as asecond control function of commanded system performance when saidmeasured prime mover speed exceeds said target speed value, whereby theload on said prime mover is adjusted to converge said prime mover speedwith said target prime mover speed.
 10. The power delivery system ofclaim 9 further comprising a clutch which is disengaged when saidvehicle is stationary, and is partially engaged during slow vehicleoperation:first predetermining means for predetermining a first fuelfunction defining desired fuel requirements for said prime mover inrelation to prime mover operating speed; second predetermining means forpredetermining a second fuel function defining desired fuel requirementsfor said prime mover in relation to command system performance; andcontrol means for controlling said fuel delivery means in accordancewith said first fuel function when said clutch is fully engaged and saidmeasured prime mover speed equals said target prime mover speed valueand for controlling said fuel delivery means in accordance with saidsecond fuel function when said clutch is less than fully engaged or saidmeasured prime mover speed does not equal said target prime mover speedvalue.
 11. The power delivery system of claim 9 further comprising:meansfor measuring the speed of the output shaft; means for generating aratio signal equal to the quotient of measured prime mover speed andmeasured output shaft speed; and means for comparing said ratio signalto a predetermined slow operation ratio, clutch slippage occurring whensaid ratio signal is greater than said slow operation ratio.
 12. Thepower delivery system of claim 10 wherein said first control functioncontrols said transmission as a function of measured system outputtorque, measured prime mover speed and said ratio signal, and saidsecond control function controls said transmission as a function ofcommanded system performance.
 13. The power delivery system of claim 12wherein said ratio control means comprises means for controlling oneside of said transmission as a function of a command system performance,measured output shaft speed and measured prime mover speed.
 14. A powerdelivery system as recited in claim 9 wherein said prime mover is aninternal combustion engine.
 15. A power delivery system as recited inclaim 9 wherein said fuel delivery means comprises a throttle.
 16. Apower delivery system as recited in claim 9 wherein said prime mover isan engine and said power driven device is an engine driven vehicle. 17.A method of controlling the operation of the engine of an engine-drivenvehicle having fuel delivery means for delivering a variable quantity offuel thereto, said engine coupled by a clutch to a continuously variabledrive ratio transmission for delivering power from said engine to anoutput shaft, said vehicle having command means for commanding a desiredoutput power or torque delivered to said oputput shaft, the drive ratioof said transmission varying as a function of commanded power or torqueto thereby cause the speed of said engine to vary, the method comprisingthe steps of:measuring the actual performance of said engine of saidengine-driven vehicle; measuring the speed of said engine; comparingmeasured engine speed to a target engine speed value; and controllingthe ratio of said transmission by controlling one side of saidtransmission as a function of the desired system performance commandedby said command means, and controlling the other side of saidtransmission as a first control function of said measured actual systemperformance when said measured engine speed equals said target enginespeed value, and as a second control function of commanded systemperformance when said measured engine speed exceed said target speedvalue, whereby the load on said engine is adjusted to converge saidengine speed with said target engine speed.
 18. The method of claim 17wherein said clutch is disengaged when said vehicle is stationary, andis partially engaged during slow vehicle operation, further comprisingthe steps of:predetermining a first fuel function defining desired fuelrequirements for said engine in relation to engine operating speed;predetermining a second fuel function defining desired fuel requirementsfor said engine in relation to command system performance; controllingsaid fuel delivery means in accordance with said first fuel functionwhen said clutch is fully engaged and said measured engine speed equalssaid target engine speed value; and controlling said fuel delivery meansin accordance with said second fuel function when said clutch is lessthan fully engaged or said measured engine speed does nto equal saidtarget engine speed value.
 19. A system for controlling the operation ofthe engine of an engine-driven vehicle having fuel delivery means fordelivering a variable quantity of fuel thereto, said engine coupled by aclutch to a continuously variable drive ratio transmission fordelivering power from said engine to an output shaft, said vehiclehaving command means for commanding a desired output power or torquedeliverd to said output shaft, the drive ratio of said transmissionvarying as a function of commanded power or torque to thereby cause thespeed of said engine to vary, the system comprising:actual systemperformance measuring means for measuring the actual performance of theoperation of said engine; speed measuring means for measuring the speedof said engine; means for comparing the measured engine speed to atarget value speed; ratio control means operatively coupled to saidcommand means and said acutal system performance measuring means forcontrolling the ratio of said transmission by controlling one side ofsaid transmission as a function of the desired system performancecommanded by said command means and controlling the other side of saidtransmission as a first control function of the measured actual systemperformance when said measured engine speed equals said target enginespeed value, and as a second control function of commanded systemperformance when said measured engine speed exceeds said target speedvalue, whereby the load on said engine is adjusted to converge saidengine speed with said target engine speed.